The present invention relates to a slide vane machine based on the geometric principle of the conchoidal of the circle, also called Pascalian Screw, with one or multiple slide vanes per working chamber in which shape and proportions of the rotating machine parts are adjusted and optimized relative to one another based on a calculation method.
Slide vane machines of the aforementioned kind are known from the U.S. Pat. Nos. 1,994,245 and DD 46,761 and have the following constructive design:
As shown in FIG. 5, in a hollow cylinder 1 with an inner contour E corresponding to the circumferential line of the Pascalian Screw which is closed off on both ends by housing end plates a rotary piston 4 in the form of a hollow circular cylinder is supported in both housing end plates in a rotatable manner and axially oriented relative to the hollow cylinder. The axis of rotation of the rotary piston corresponds to the center M' of the central circle K' of the Pascalian Screw represented in drawings 1 and 2. From this it results that the rotary piston with the radius r.sub.Kr with its outer surface approaches the inner surface of the hollow cylinder at the approximation point 21(-xr.sub.Kr Yo)=E180.degree. to form a narrow slot.
The control shaft 8, embodied as a crankshaft, with the radius r of the base circle K is guided through the hollow rotary piston and through central openings in its end faces and is also rotatably supported within the two housing end plates, whereby the crankshaft axis corresponds to the center M of the base circle K, the crankshaft axis being arranged parallel to the axis of the rotary piston.
On the crankshaft pins of the control shaft one or two slide vanes 6 with their half length L/2 are rotatably supported. They correspond functionally to slide vane wing pairs 6f staggered by 180.degree. which are connected to form a rigid slide vane by stays 6s connected to one or two central bearing sleeves 61. Their center P corresponds to the axis of the crankshaft pins, and during the rotation of the control shaft, they follow the base circle K. The slide vanes are additionally guided in diametrically opposed longitudinal grooves 5 of the rotary piston
In machines with one slide vane per working chamber the masses due to inertia of the crankshaft pins revolving on the base circle K and of the slide vane must be compensated by two counterweights at the control shaft.
The use of more than one slide vane per working chamber allows for the compensation of the forces of inertia without counterweights, when the slide vanes have a longitudinal and transverse axial symmetry and the bifurcation of the stays 6s is arranged with the bearing sleeve 61 is arranged such that they do not overlap one another.
Since the axis of rotation of the rotary piston is at M'.sub.(xo yo) =P.sub.(xo yo), the central circle K' contacts the base circle K at P.sub.(x2r yo) =S.sub.(x2r yo), r' corresponds to 2r, and the points P and S have the same trajectory velocity and direction of rotation, the central outer toothing of the control shaft 8 with the partial circle K.sub.t may engage a central inner toothing (see, for example, the representation of FIG. 9) within one rotary piston end face with the partial circle K.sub.t '.
The ratio of the number of teeth of the outer toothing 10 to the number of teeth of the inner toothing 11 is 1:2. The toothing engagement of the control shaft with the rotary piston ensures the exact synchronous rotation of these two machine parts at the ratio 2:1 and compensates bending forces acting on the slide vanes.
The machine geometry according to the Pascalian Screw also allows for constructions without toothing engagement of the control shaft 8 with the rotary piston 4, because the course of movement of the slide vanes effects a force-controlled rotation of the rotary piston and the control shaft in the aforementioned ratio 1:2.
However, due to the torque acting on the control shaft, an additional bending moment acts on the slide vanes. The bending of the slide vanes may lead to canting within the longitudinal grooves 5.
When synchronously rotating control shaft and rotary piston, the mass points of the slide vanes which are centrally located on the axis within the point P move with the number of revolutions of the control shaft along the base circle K with simultaneous rotation of the slide vanes about their axis in the point P with the number of revolutions of the rotary piston (see FIG. 5). The rounded side edges 22 of the slide vanes follow the circumferential line E of the Pascalian Screw and may thus be guided contact-free along the inner surface of the hollow cylinder 1 with the same contour.
Simultaneously, the slide vanes perform a periodic pushing movement relative to the rotary piston side wall which per revolution in each direction reaches once the maximum displacement NE.sub.max =2D at the position 0.degree..
In the patent DD 46 761 a mounting of only one slide vane per working chamber for a working principle which differs from the object of the present invention is disclosed.
The gas outflow takes place via channels within the rotary piston and the desired high inner compression is accomplished by employing the rotary piston 4 as the slide vane valve.
In the text and in the drawings the control shaft 8 is represented as a crankshaft. The problem of the design and lubrication of the slide vane bearing is not addressed.
In the text and the drawings of U.S. Pat. No. 1,994,245, the control shaft is also represented as a crankshaft with a lubricant channel that is guided through the crankshaft pins and the webs.
The constructive principle upon which the above invention is based discloses only the use of two slide vanes per working chamber. The slide vanes are comprised of two halves which are divided at the respective bearing.
The aforementioned prior art devices have the following disadvantages:
The control shaft is in the form of a crankshaft. Accordingly, the slide vanes are provided with the greatest possible constructive space within the rotary piston; however, the crankshaft concept has the following severe disadvantages:
Due to the low eccentricity the crankshaft may be manufactured in a divided form with known screw and plug-in connections which allow the use of undivided needle bearings only under high expenditures with a strongly reduced bending and torsional stiffness. Accordingly, as already suggested in U.S. Pat. No. 1,994,245, the control shaft is thus embodied as a undivided crankshaft with a lubricant channel within the crankshaft pins and webs. The slide vanes must then be provided with divided friction bearings.
However, with this embodiment a new problem arises: The boring of the lubricant channel. Due to the manufacturing technology it must be assembled from a plurality of partial borings extending through the crankshaft, the crankshaft pin, and the crankshaft webs. Due to the low eccentricity of the control shaft the through boring of the crankshaft pin for a higher number of slide vanes is very difficult.
The control shaft, due to the machine geometry, may be supported with its ends only once at the housing end plates. This results, for a greater constructive length which is unavoidable when using a greater number of slide vanes and which, due to the favorable shape of the slide vane machine, is also desirable, in the embodiment as a crankshaft to a reduced bending and torsion stiffness because of the smaller crankshaft pin cross-sections.
Accordingly, the load capacity of the control shaft and the precision of the slot between the slide vane side edges 22 and the hollow cylinder inner surface area is lowered parallel to a respective reduction of the efficiency.
The available area for the positioning of the slide vane bearings on the control shaft is twofold reduced with the embodiment as a crankshaft:
Due to the smaller diameter of the crankshaft pin the bearing diameters are small.
The maximum total width of the bearing is determined by the control shaft section within the hollow cylinder. With an increasing number of slide vanes the bearing width that is available for the individual slide vane is thus reduced. It is further reduced due to the necessary formation of the crankshaft webs.
The aforementioned disadvantages may be avoided when the control shaft is embodied as an eccentric shaft:
The manufacture as a divided shaft is substantially simpler. As can be seen from drawing FIG. 5, the individual eccentric segments 8s with a respective angular staggering are positioned on a continuous central guide shaft 9. The slide vanes may also be provided with undivided needle bearings.
The lubricant channel may be embodied as a central longitudinal slotted boring 16 from which extend radial bores within the control shaft eccentrics to the slide vane bearings.
In addition to the production-technological advantages a functional advantage in the form of an increase in the lubricant pressure due to the centrifugal acceleration within the radial eccentric borings is achieved.
The bending and torsional stiffness is increased due to the axially overlapping eccentric segments 8s.
The diameter and the width of the slide vane bearings are greater.
In the embodiment as an eccentric shaft the control shaft cross-section is greater and, accordingly, the rotary piston diameter D.sub.Kr for a constant maximum displacement NE.sub.max is also increased.
The technical applicability is thus limited because with an increasing diameter of the control shaft the ratio of the maximum displacement NE.sub.max the rotary piston diameter D.sub.Kr and thus the specific displacement volume is reduced. The displacement of the proportion of L and D.sub.Kr to NE.sub.max is represented in drawing FIG. 3.
Simultaneously, the toothing diameter of the control shaft in relation to the slide vane length is reduced. The smaller toothing is loaded by an increased torque.
A further disadvantage of the prior art is the low number of slide vanes per working chamber:
In analogy to the working principle of the multi-cellular machine a number of slide vanes as high as possible per working chamber should be realized in order to accomplish comparable pulsation and inner compression values.
FIG. 9 shows an exemplary embodiment of a slide vane machine according to the principles of the present invention. The housing 1 has end walls 2, 3 in which a rotary piston in the form of a hollow circular cylinder 4 is rotatably supported by bearings 15. A control shaft 8 with sections 8s is connected to a central guide shaft 9 and is supported in bearings 13 within the end walls 2, 3. The outer toothing 10 of the control shaft 8 cooperates with the inner toothing 11 of the rotary piston 4. The slide vanes are wing pairs 6f that are connected via a center stay 6g with central bearing sleeves 61 for supporting the slide vane bearings 14 to form a fixed slide vane.
Slide vane machines with one or two slide vanes are technically of no importance because of pulsation values that are too great and achievable maximum inner compression rates that are too low.
The slide vane machine of the aforementioned U.S. patent within the represented proportion of D.sub.Kr to NE.sub.max and the design with two slide vanes per working chamber may not be expanded to a greater number.
From the prior art it may not be deduced whether the construction principle may be expanded to a number of slide vanes comparable to multi-cellular machines because of the additionally required constructive space for the use of an eccentric shaft.
It is therefore an object of the present invention to realize the embodiment of the control shaft as an eccentric shaft while simultaneously employing a greater number of slide vanes per working chamber with the aid of a calculation method or rule for slide vane machines of the aforementioned kind and, on the other hand, to determine the shape and proportionality of control shaft with the outer toothing 10, slide vanes, and rotary piston with the inner toothing 11 with optimal use of the constructive space such that the maximum displacement NE.sub.max reaches the greatest ratio possible with respect to the diameter D.sub.Kr of the rotary piston.